Method and apparatus for cooling an airfoil

BACKGROUND OF THE INVENTION

1. Technical Field

This invention relates to gas turbine engines in general, and to methods and apparatus for cooling a substrate exposed to high temperature gas in particular.

2. Background Information

Efficiency is a primary concern in the design of any gas turbine engine. Historically, one of the principle techniques for increasing efficiency has been to increase the core gas path temperatures within the engine. Core gas refers to air worked within the compressor that is mixed with fuel and combusted within the combustor. The increased gas path temperatures have been accommodated by using internally cooled components made from high temperature capacity alloys. Turbine stator vanes and blades, for example, are typically cooled using compressor air worked to a higher pressure, but still at a lower temperature than that of the core gas flow passing by the blade or vane. The higher pressure provides the energy necessary to push the air through the component. A significant percentage of the work imparted to the air bled from the compressor, however, is lost during the cooling process. The lost work does not add to the thrust of the engine and therefore negatively effects the overall efficiency of the engine. A person of skill in the art will recognize, therefore, that there is a tension between the efficiency gained from higher core gas path temperatures and the concomitant need to cool turbine components and the efficiency lost from bleeding air to perform that cooling.

There is, accordingly, great value in maximizing the cooling effectiveness of whatever cooling air is used. Prior art coolable airfoils typically include a plurality of internal cavities, which are supplied with cooling air. The cooling air passes through the wall of the airfoil (or the platform) and transfers thermal energy away from the airfoil in the process. The manner in which the cooling air passes through the airfoil wall is critical to the efficiency of the process. In some instances, cooling air is passed through straight or diffused cooling apertures to convectively cool the wall and establish an external film of cooling air. A minimal pressure drop is typically required across these type cooling apertures to minimize the amount of cooling air that is immediately lost to the free-stream hot core gas passing by the airfoil. The minimal pressure drop is usually produced through a plurality of cavities within the airfoil connected by a plurality of metering holes. Too small a pressure drop across the airfoil wall can result in undesirable hot core gas in-flow. In all cases, the minimal dwell time in the cooling aperture as well as the size of the cooling aperture make this type of convective cooling relatively inefficient.

Some airfoils convectively cool by passing cooling air through passages disposed within a wall or platform. Typically, those passages extend a significant distance within the wall or platform along a substantially straight line. There are several potential problems with this type of cooling scheme. First, the heat transfer rate between the passage walls and the cooling air decreases markedly as a function of distance traveled within the passage. As a result, cooling air flow adequately cooling the beginning of the passage may not adequately cool the end of the passage. If the cooling air flow is increased to provide adequate cooling at the end of the passage, the beginning of the passage may be excessively cooled, consequently wasting cooling air. Second, the thermal profile of an airfoil is typically non-uniform and will contain regions exposed to a greater or lesser thermal load. The prior art internal cooling passages extending a significant distance within an airfoil wall or a platform typically span one or more regions having disparate thermal loads. Similar to the situation described above, providing a cooling flow adequate to cool the region with the greatest thermal load can result in other regions along the passage being excessively cooled.

What is needed, therefore, is a method and apparatus for cooling a substrate within gas turbine engine that adequately cools the substrate using a minimal amount of cooling air and one that provides heat transfer where it is needed.

DISCLOSURE OF THE INVENTION

It is, therefore, an object to provide a method and an apparatus for cooling a wall within a gas turbine engine for removing more cooling potential from cooling air passed through the wall than is possible using most conventional methods and apparatus.

It is another object to provide a method for cooling a wall within a gas turbine engine that can produce a cooling profile that substantially matches the thermal profile of the wall, and an apparatus that can be used for the same.

According to the present invention, an apparatus and a method for cooling a wall for use in a gas turbine engine is provided that includes a cooling microcircuit. The cooling microcircuit, which can be disposed within the wall of a component such as a stator vane or a rotor blade, includes passage having a plurality of segments connected in series by one or more chambers. An inlet aperture connects the passage to one side of the wall. An exit aperture connects the passage to the opposite side of the wall. Cooling air on the inlet aperture side of the wall enters the passage through the inlet aperture and exits through the exit aperture.

The present cooling apparatus and method for cooling a wall provides significantly increased cooling effectiveness over prior art cooling schemes. One of the ways the present apparatus and method provides increased cooling effectiveness is by increasing the heat transfer coefficient per unit flow within a cooling passage. The transfer of thermal energy between the wall containing the passage and the cooling air is directly related to the heat transfer coefficient within the passage for a given flow. A velocity profile of fluid flow adjacent each wall of a passage is characterized by an initial hydrodynamic entrance region and a subsequent fully developed region as can be seen in FIG.<b>7</b>. In the entrance region, a fluid flow boundary layer develops adjacent the walls of the passage, starting at zero thickness at the passage entrance and eventually becoming a constant thickness at some position downstream within the passage. The change to constant thickness marks the beginning of the fully developed flow region. The heat transfer coefficient is at a maximum when the boundary layer thickness is equal to zero, decays as the boundary layer thickness increases, and becomes constant when the boundary layer becomes constant. Hence, for a given flow the average heat transfer coefficient in the entrance region is higher than the heat transfer coefficient in the fully developed region. The present apparatus and method increases the percentage of flow in a passage characterized by entrance region effects by providing a plurality of short length segments connected by chambers. Fluid entering a chamber diffuses and decreases in velocity. Fluid exiting a chamber is characterized by entrance region effects and consequent increased local heat transfer coefficients. The average heat transfer coefficient per unit flow of the relatively short segments of the present apparatus and method is consequently higher than that available in all similar prior art cooling schemes of which we are aware.

Another way the present invention provides an increased cooling effectiveness also involves the short length segment between chambers. The relationship between the beat transfer rate and the heat transfer coefficient in given length of passage can be mathematically described as follows:

<i>q=h</i>_c<i>A</i>_s<i>ΔT</i>_lm  (Eqn. 1)

where:

q=heat transfer rate between the passage and the fluid

h_c=heat transfer coefficient of the passage

A_s=passage surface area =P×L=Passage perimeter×length

ΔT_lm=log mean temperature difference

The above equation illustrates the direct relationship between the heat transfer rate and the heat transfer coefficient, as well the relationship between the heat transfer rate and the difference in temperature between the passage surface temperature and the inlet and exit fluid temperatures passing through a length of passage (i.e., ΔT_lm). In particular, if the passage surface temperature is held constant (a reasonable assumption for a given length of passage within an airfoil, for example) the temperature difference between the passage surface and the fluid decays exponentially as a function of distance traveled through the passage. The consequent exponential decay of the heat transfer rate is particularly significant in the fully developed region where the heat transfer coefficient is constant and the heat transfer rate is dependent on the difference in temperature. The present apparatus and method use relatively short length segments disposed between chambers. As stated above, cooling airflow passing through a portion of each segment is characterized by an entrance region velocity profile and the remainder is characterized by a fully developed velocity profile under normal operating conditions. In all embodiments of the present apparatus and method, the segment length between chambers is short to minimize the effect of the exponentially decaying heat transfer rate attributable to temperature difference, particularly in the fully developed region.

In some embodiments, the present apparatus and method includes a number of segments successively shorter in length. The longest of the successively shorter segments is positioned adjacent the inlet aperture where the temperature difference between the fluid temperature and the passage wall is greatest, and the shortest of the successively shorter segments is positioned adjacent the exit aperture where the temperature difference between the fluid temperature and the passage wall is smallest. Successively decreasing the length of the segments within the passage helps to offset the decrease in ΔT_lm, in each successive segment. For explanation sake, consider a plurality of same length segments, connected to one another in series. The average ΔT_lmof each successive segment will decrease because the cooling air increases in temperature as it travels through each segment. The average heat transfer rate, which is directly related to the ΔT_lm, consequently decreases in each successive segment. Cooling air traveling through a plurality of successively shorter segments will also increase in temperature as it passes through successive segments. The amount that the ΔT_lmdecreases per segment, however, is less in successively shorter segments (vs. equal length segments) because the length of the segment where the exponential temperature decay occurs is shorter. Hence, decreasing segment lengths positively influence the heat transfer rate by decreasing the influence of the exponential decaying temperature difference.

Another way the present invention provides an increased cooling effectiveness is by utilizing cooling air pressure difference in a manner that optimizes heat transfer within the passage. Convective heat transfer is a function of the Reynolds number and therefore the Mach number (i.e., velocity) of the cooling air traveling through a segment of the passage. In one embodiment of the present apparatus and method, cooling air is maintained at substantially the same Mach number in each segment by maintaining the substantially the same ratio of chamber pressures across each segment. The preferred manner for maintaining substantially the same ratio of chamber pressures across each segment is by altering the cross-sectional area of each successive segment within the passage.

The small size of the present cooling apparatus also provides advantages over many prior art cooling schemes. The thermal profile of most blades or vanes is typically non-uniform along its span and/or width. If the thermal profile is reduced to a plurality of regions however, and if the regions are small enough, each region can be considered as having a uniform heat flux. The non-uniform profile can, therefore, be described as a plurality of regions, each having a substantially uniform heat flux albeit different in magnitude. The present cooling microcircuits can be sized to fit within most of those regions of uniform heat flux. Consequently, an embodiment of the present microcircuit can be tuned and deployed to offset a particular magnitude heat flux present in a particular region. A blade or vane having a non-uniform thermal profile, for example, can be efficiently cooled with the present invention by positioning one or more microcircuits at particular locations within the blade or vane wall, and matching the cooling capacity of the microcircuit(s) to the local heat flux. As a result, excessive cooling is decreased and the cooling efficiency is increased.

The size of the present cooling microcircuit also provides cooling passage compartmentalization. Some conventional cooling passages include a long passage volume connected to the core gas side of the wall by a plurality of exit apertures. In the event a section of the passage is burned through, it is possible for a significant portion of the passage to be exposed to hot core gas in-flow through the plurality of exit apertures. The present apparatus and method limits the potential for hot core gas in-flow by preferably utilizing only one exit aperture per passage. In the event hot core gas in-flow does occur, the present passages are limited in area, consequently limiting the area potentially exposed to undesirable hot core gas.

These and other objects, features and advantages of the present invention will become apparent in light of the detailed description of the best mode embodiment thereof, as illustrated in the accompanying drawings.

BRIEF DESCRIPTION OF THE DRAWINGS

FIG. 1 is a diagrammatic view of a gas turbine engine.

FIG. 2 is a diagrammatic view of a rotor blade having a plurality of the present invention cooling microcircuits disposed in a wall.

FIG. 3 is an enlarged diagrammatic perspective view of an embodiment of the present invention cooling microcircuit.

FIG. 4 is a diagrammatic planar view of the present invention cooling microcircuit embodiment shown in FIG. <b>3</b>.

FIG. 5 is a diagrammatic view of an embodiment of the present invention cooling microcircuit.

FIG. 6 is a diagrammatic view of an embodiment of the present invention cooling microcircuit.

FIG. 7 is a fluid flow velocity profile chart illustrating a velocity profile having an entrance region followed by a fully developed region.

DETAILED DESCRIPTION OF THE INVENTION

Referring to FIGS. 1-3, a cooling microcircuit <b>10</b> is disposed within a wall <b>12</b> exposed to hot core gas within a gas turbine engine <b>14</b>. Cooling air is typically present on one side of the wall <b>12</b> and hot core gas is present on the opposite side of the wall <b>12</b> during operating conditions. Potential applications of the present invention microcircuit <b>10</b> include, but are not limited to, combustors <b>16</b> and combustor liners <b>18</b>, blade outer air seals <b>20</b>, turbine exhaust liners <b>22</b>, augmentor liners <b>24</b>, nozzles <b>26</b>, stator vanes <b>28</b>, and rotor blades <b>30</b>. For purposes of providing a detailed description, the present microcircuit <b>10</b> will be described in the context of a rotor blade <b>30</b> application. FIG. 2 shows the microcircuit <b>10</b> disposed in an airfoil portion <b>32</b> of a turbine rotor blade <b>30</b>, although the microcircuit <b>10</b> may also be disposed in the platform portion <b>34</b>.

Referring to FIG. 3, a cooling microcircuit <b>10</b> is shown disposed in a wall <b>12</b> between a first surface <b>33</b> and a second surface <b>35</b> of the wall <b>12</b>. Each microcircuit <b>10</b> embodiment includes a passage <b>11</b> consisting of a plurality of segments <b>36</b> connected in series by one or more chambers <b>38</b>, an inlet aperture <b>40</b>, and an exit aperture <b>42</b>. Each chamber <b>38</b> has a cross-sectional flow area greater than the cross-sectional flow area of a segment <b>36</b>. The cross-sectional flow of each chamber is great enough to cause cooling air flow exiting a segment <b>36</b> to diffuse and decrease in velocity, and great enough to cause flow exiting a chamber to nozzle and increase in velocity. The changes in cooling air velocity and the resultant changes in cooling air pressure across a segment <b>36</b> illustrate how the cooling air is metered within the segment <b>36</b>. The inlet aperture <b>40</b> connects the passage <b>11</b> to one side of the wall <b>12</b>. The exit aperture <b>42</b> connects the passage <b>11</b> to the opposite side of the wall <b>12</b>. The inlet and exit apertures <b>40</b>,<b>42</b> may be disposed in a segment <b>36</b> or in a chamber <b>38</b>. Cooling air on the inlet aperture side of the wall <b>12</b> enters the passage <b>11</b> through the inlet aperture <b>40</b> and exits through the exit aperture <b>42</b>.

Each cooling microcircuit embodiment can occupy a wall surface area as great as 0.1 square inches (64.5 mm²). It is more common, however, for a microcircuit <b>10</b> to occupy a wall surface area less than 0.06 square inches (38.7 mm²), and the wall surface of preferred embodiments typically occupy a wall surface area closer to 0.01 square inches (6.45 mm²). Passage segment size will vary depending upon the application, but in most embodiments the cross-sectional area of the segment <b>36</b> is less than 0.001 square inches (0.6 mm²). The most preferred segment embodiments have a cross-sectional area between 0.0001 and 0.0006 square inches (0.064 mm²and 0.403 mm²) with a substantially rectangular shape. For purposes of this disclosure, segment <b>36</b> (or chamber <b>38</b>) cross-sectional area shall be defined as a cross-section taken along a plane substantially perpendicular to the direction of cooling airflow through the segment <b>36</b> (or chamber <b>38</b>).

Referring to FIGS. 3 and 4, in one embodiment of the present microcircuit <b>10</b> the passage <b>11</b> includes a series of segments <b>36</b> connected by chambers <b>38</b>, disposed in a spiral-like arrangement. The example of this embodiment shown in FIGS. 3 and 4, includes four segments <b>36</b> and five chambers <b>38</b>. The first segment <b>44</b> and the third segment <b>46</b> are substantially parallel one another and connected via a second segment <b>48</b> that extends substantially perpendicular to the first and third segments <b>44</b>,<b>46</b>. The fourth segment <b>50</b> extends into the area boxed by the first, second, and third segments <b>44</b>,<b>48</b>,<b>46</b>, giving the microcircuit <b>10</b> its spiral-like configuration. The first chamber <b>52</b> is attached to one end of the first segment <b>44</b>. The inlet aperture <b>40</b> is disposed in the first chamber <b>52</b>, connecting the passage <b>11</b> to one side of the wall <b>12</b>. The second chamber <b>54</b> connects the first and second segments <b>44</b>,<b>48</b>, the third chamber <b>56</b> connects the second and third segments <b>48</b>,<b>46</b>, and the fourth chamber <b>58</b> connects the third and fourth segments <b>46</b>,<b>50</b>. The fifth chamber <b>60</b> is attached to the end of the fourth segment <b>50</b>. The exit aperture <b>42</b> is disposed in the fifth chamber <b>60</b>, connecting the passage <b>11</b> to the opposite side of the wall <b>12</b>.

FIG. 5 shows an alternative embodiment of the present microcircuit <b>10</b> in which the passage includes four chambers <b>38</b> and three segments <b>36</b> arranged in a substantially linear configuration. Passages <b>11</b> can assume a variety of configurations of segments <b>36</b> in series connected by chambers <b>38</b>, and are not limited to the examples given here for explanation sake.

Referring to FIG. 6, in some embodiments the passage <b>11</b> includes a number of segments <b>36</b> successively shorter in length (L₁&gt;L₂&gt;L₃) connected by chambers <b>38</b>. The longest of the successively shorter segments <b>36</b> communicates with the inlet aperture <b>40</b>. At the inlet aperture <b>40</b>, the temperature difference between the fluid temperature and the passage wall is greatest. The shortest of the successively shorter segments <b>36</b> communicates with the exit aperture <b>42</b>. At the exit aperture <b>42</b>, the temperature difference between the fluid temperature and the passage wall is smallest. Successively decreasing the length of the segments <b>36</b> within the passage <b>11</b> helps to offset the decrease in ΔT_lmin each successive segment <b>36</b>. The successively decreased segment lengths positively influence the heat transfer rate by decreasing the influence of the exponential decaying temperature difference.

Referring to FIGS. 3 and 4, in some embodiments each successive segment <b>36</b> has a cross-sectional area greater than the prior or “upstream” segment <b>36</b> (e.g., the cross-sectional area for the second segment is greater than the cross-sectional area for the first segment). The increase in segment cross-sectional area (A_Sn) can, for example, be accomplished by holding the height (H) of the segments <b>36</b> constant and increasing the width (W_n) of the successive segments <b>36</b> (A_S1&lt;A_S2&lt;A_S3&lt;A_S4, where A_Sn=W_n×H). The change in cross-sectional area per segment <b>36</b> is chosen to create a substantially constant chamber pressure (P_Cn) ratio across each segment <b>36</b> for a given set of operating conditions (e.g., P_C1/P_C2≈P_C2/P_C3). The substantially constant chamber pressure ratio across each segment <b>36</b> produces a cooling air velocity in each segment <b>36</b> that substantially equals the cooling air velocity in each other segment <b>36</b>. As a result, the cooling air is metered substantially equally across each segment <b>36</b> rather than just across the inlet and exit apertures <b>40</b>,<b>42</b>. As stated above, convective heat transfer is a function of the Reynolds number and therefore the Mach number of the cooling air traveling within a segment <b>36</b>. The ability of the present microcircuit <b>10</b> to provide a cooling air velocity substantially equal in each segment <b>36</b> enables the microcircuit <b>10</b> to provide an optimum Mach number for a given set of operating conditions and therefore an optimum heat transfer for those operating conditions.

Under typical operating conditions within the turbine section of a gas turbine engine, the cooling air Mach number within the microcircuit will likely be in the vicinity of 0.3. With a Mach number in that vicinity, the entrance region within a typical segment <b>36</b> will likely extend somewhere between five and fifty diameters (diameter=the segment hydraulic diameter). Obviously, the length of the segment <b>36</b> will dictate what segment length percentage is characterized by velocity profile entrance region effects; e.g., a shorter segment will have an increased percentage of its length characterized by velocity profile entrance effects. Preferably, a segment <b>36</b> within the present microcircuit <b>10</b> will have at have least fifty percentage of its length devoted to entrance region effects.

For any given set of operating conditions, each of the above described microcircuit <b>10</b> embodiments will provide a particular heat transfer performance. It may be advantageous, therefore, to use more than one embodiment of the present microcircuit in those applications where the thermal profile of the wall to be cooled is non-uniform. The microcircuits <b>10</b> can be distributed to match and offset the non-uniform thermal profile of the wall <b>12</b> and thereby increasing the cooling efficiency of the wall <b>12</b>.

Although this invention has been shown and described with respect to the detailed embodiments thereof, it will be understood by those skilled in the art that various changes in form and detail thereof may be made without departing from the spirit and the scope of the invention. For example, the detailed description above gives the preferred embodiment wherein the chamber pressure ratio across each segment in a passage is substantially equal to the chamber pressure ratio across other segments within the passage. In some instances, however, it may advantageous to vary the chamber pressure ratios across segments within a passage to suit the cooling application at hand.